The present invention relates to steam turbines. In particular, the invention relates to the configuration of the turbine blades for a steam turbine.
With recent turbines, there has been a tendency to use longer blades in the final turbine stage and in the turbine stages upstream of the final stage to economise on fuel and operate more efficiently.
For example, FIG. 10 shows a 700,000 kW-output class steam turbine in which long blades have been adopted in the final turbine stage and the turbine stages upstream of the final turbine stage. This is an axial flow type turbine in which multiple stages 5 are located serially in the turbine-driving steam flow along the axial direction of turbine shaft 2 that is housed in turbine casing 1. Each stage 5 comprises a set of fixed turbine nozzle blades 3, and a downstream adjacent set of turbine moving blades 4.
The turbine nozzle blades 3 of each stage are aligned in the circumferential direction around the turbine shaft 2 with their outer ends supported by an outer diaphragm 6, which is fixed in the turbine casing 1, and their inner ends supported by an inner diaphragm 7 adjacent the turbine shaft 2. A seal 7a carried by the inner diaphragm 7 seals inner diaphragm 7 to rotating shaft 2.
The turbine moving blades 4 of each stage are circumferentially aligned around turbine shaft 2, adjacent and downstream of the turbine nozzle blades 3 of that stage. Each turbine moving blade extends radially from the shaft 2 and has a blade embedded portion 8 embedded in the shaft 2, a blade effective portion 9 from root to tip and a blade tip connecting portion 10. The blade effective portion 9 is the part of the blade that does the actual work (generates rotational torque) when the turbine driving steam passes through the turbine moving blades.
The turbine moving blades 4 are provided with intermediate connectors 11 in the intermediate parts of the blade effective portions 9, which serve to stabilize the effective portions 9 of the entire set of blades. The intermediate connectors 11 comprise, as shown in FIG. 11, bosses 11a and 11b on the respective backs (xe2x80x9csuction sidexe2x80x9d or xe2x80x9csuction surfacexe2x80x9d as it is commonly called), 9c and 9d, and bellies (xe2x80x9cpressure sidexe2x80x9d or xe2x80x9cpressure surfacexe2x80x9d as it is commonly called), 9e and 9f, of one blade effective portion 9a and the adjacent blade effective portion 9b. A linking sleeve 11c pivotally interconnects bosses 11a and 11b via lugs (not shown) provided at both ends of bosses 11a and 11b. Thus, vibration of the intermediate portions, induced by such factors as fluctuations over time of the jet force of the turbine driving steam flowing from the turbine nozzle blades 3, and turbine shaft vibration, is suppressed to a low level.
The tips of turbine moving blades 4 are stabilized by blade tip connectors 10 which are formed, for example, as so-called xe2x80x9csnubber typexe2x80x9d plate-shaped extension pieces 10a and 10b integrally cut from the blade effective portion 9, as shown in FIG. 12. During operation, blade tip vibration is suppressed using the mutual contact friction of the extension pieces 10a and 10b. 
The above-described arrangement of intermediate connectors 11 and blade tip connectors 10 provides effective countermeasures against vibration induced by such factors as variation over time of the turbine driving steam jet force, in turbines having long blades. However, in a prior art steam turbines (shown in FIG. 10), with long blades in which the blade effective portions 9 of the turbine moving blades 4 exceed 1 m, many other problems arise because of the blade length. One of these is that, during operation, the throat-pitch ratio (S/T) varies as a consequence of deformation of the blade warp configuration due to centrifugal force, resulting in a reduction of aerodynamic efficiency.
Attempts have been made in the prior art to address this problem by adopting the so-called xe2x80x9csimplified three-dimensional blade design methodxe2x80x9d. In this method, the cross-sectional shape of the turbine moving blade is varied to correspond to the fact that the equivalent velocity diagram had been largely changed in the height direction of the passage. However, if the turbine moving blades 4 of the steam turbine are long, as shown in FIG. 13, the inlet flow angle of the turbine driving steam relative to the turbine blade will vary greatly along the blade effective portion 9 from the blade root to the blade mean diameter (pitch circle diameter), to the blade tip.
In FIG. 13, xcex1 indicates the inlet flow angle of the turbine driving steam to the turbine moving blade 4, BV the turbine driving steam inlet flow speed vector flowing into the turbine moving blade 4, SV the turbine driving steam outlet flow speed vector flowing out of the turbine nozzle blades (not shown) and U the peripheral speed, respectively. Also, the subscripts R, P and T indicate the respective blade root, blade mean diameter (pitch circle diameter) and blade tip position.
In this case, there is a requirement to modify the blade cross-sectional shapes at each of the blade root, the blade mean diameter and the blade tip positions of the blade effective portion 9 to correspond to the turbine driving steam inlet flow angles xcex1R, xcex1P and xcex1T at each position. However, as a prerequisite for that, first there is a requirement to find turbine driving steam inlet flow speed vectors BVR, BVP and BVT at each position.
Turbine driving steam inlet flow speed vectors BVR, BVP and BVT at each position can be found from equivalent velocity diagrams composed of outlet flow speeds SVR, SVP and SVT of the turbine driving steam flowing out from the blade root, the blade mean diameter and the blade tip positions of the turbine nozzle blades, and the circumferential speed vector (the turbine shaft circumferential speed component) determined by the radius and angular rotational speed at each position (the angular rotational speed of course being constant, independent of radial position).
For turbine driving steam inlet flow speed vectors BVR, BVP and BVT at the various positions found from equivalent velocity diagrams, the inlet flow angles can vary. For example, the inlet flow angle xcex1T at the blade root typically is in the range of about 30xc2x0 to about 50xc2x0 while the inlet flow angle xcex1T at the blade tip typically is in the range of about 140xc2x0 to about 170xc2x0, and their angular difference may be a maximum of about 140xc2x0. This large angular difference is due to the fact that the radial position of the blade tip (measured from the turbine shaft axis of rotation) is at least twice that of the blade root, and, proportionally, the circumferential speed component at the blade tip is at least twice that at the blade root.
If the turbine moving blade is not modified to compensate for this large variation in the inlet flow angle in the radial direction, aerodynamic loss will markedly increase. Therefore, prior art steam turbines were modified by varying the twist angle of the blade cross-section to conform it to the turbine driving steam inlet flow angles xcex1R, xcex1P and xcex1T at the various positions on the blade effective portion 9; and, moreover, the blade cross-sectional shape close to the leading edge was modified in the direction of the inlet flow speed vector.
FIG. 14 is a drawing of a circumferential direction cross-section at any height of the turbine moving blade row, developed on a plane, and shows the configuration of the turbine moving blade steam passage. S is the throat, and indicates the width of the narrowest part in the inter-blade steam passage formed between the back of one blade and the belly of the next turbine moving blade. T is the pitch, that is the gap between turbine moving blades in the circumferential direction. The throatxc2x7pitch ratio (S/T) is an aerodynamic design parameter that does not depend on the size of the steam turbine, and corresponds to the outlet flow angle of the turbine moving blades. In other words, if the throat-pitch ratio (S/T) is increased, the turbine moving blade outlet flow angle, which is defined by taking the circumferential direction as zero, becomes larger and, when the blade outlet flow speed is taken as constant, the axial flow speed component becomes greater and the flow rate of this cross-section increases. Conversely, if the throatxc2x7pitch ratio (S/T) is decreased, the turbine moving blade outlet flow angle becomes smaller, and the flow rate of this cross-section decreases. The definition of the throat-pitch ratio (S/T) is the same for the turbine nozzle blades also.
In long-blade stages, such as the turbine final stage, the pressure difference between the inner wall side (blade root) and the outer wall side (blade tip), due to the tangential velocity component produced by the turbine nozzle blades, becomes greater. In the design of long blade stages, it is necessary to adopt a throat-pitch ratio (S/T) distribution that takes account of this pressure difference.
FIG. 15 is an example of the turbine moving blade throat-pitch ratio (S/T) distribution normally adopted in prior art designs. In the prior art xe2x80x9csimplified three-dimensional design method,xe2x80x9d because it was difficult accurately to estimate the three-dimensional loss of each blade cross-section, designs were produced so that the flow rate distribution per unit annular area in the radial direction became approximately constant for both turbine nozzle blades and turbine moving blades. For the turbine moving blade, instead of the exit static pressure distribution being approximately constant, the flow speed increased on the outer wall side where the entry static pressure was high. Therefore, a design was adopted in which, in addition to reducing the axial flow speed by reducing the throatxc2x7pitch ratio (S/T) on the outer wall side, the axial flow speed was increased by increasing throatxc2x7pitch ratio (S/T) on the inner wall side where, conversely, the entry static pressure is low and the turbine moving blade outlet flow speed is low. Thus the radial direction flow distribution became approximately uniform.
With prior art turbine moving blades designed in this way, there are no problems when the blade height is low. However, with long blades exceeding 1 m in blade height, there is the problem that it is difficult sufficiently to ensure a pressure difference between the inlet and outlet of the blade root cross-section of the turbine moving blades that is commensurate with the relative pressure drop of the entry static pressure. This could lead to reduced performance. At the same time, by passing the same degree of flow rate both at the blade root cross-section and at other cross-sections, there is also the problem that the aerodynamic performance of the turbine stage as a whole is reduced.
FIG. 16 shows the throatxc2x7pitch ratio (S/T) distribution of a prior art turbine nozzle blade. With a turbine nozzle blade, in contrast to a turbine moving blade, as opposed to the entry total pressure being approximately uniform, the exit static pressure distribution has a distribution that increases from inside to outside. With the prior art xe2x80x9csimplified three-dimensional design method,xe2x80x9d because it was difficult to forecast the loss distribution in the radial direction, it was taken as a premise that the flow distribution in the radial direction was uniform. For this reason, the throatxc2x7pitch ratio (S/T) distribution shown in FIG. 16, which increases continuously from blade root to blade tip, was adopted.
The problem with the distribution in FIG. 16 is that, due to the outlet flow angle at the blade root becoming smaller, the loss in this part increases. Also, there is the problem that, with the blade tip being close to the wall surface, loss is increased by secondary flow turbulence occurring in the corner between the wall surface and the turbine nozzle blade. Because the same degree of flow rate as in other blade regions flows through this region too, the aerodynamic performance of the turbine stage as a whole decreases.
FIG. 17 shows the radial direction distribution of aerodynamic loss in prior art turbine nozzle blades. At the blade root side, through reducing the outlet flow angle by making the throatxc2x7pitch ratio (S/T) smaller, a vicious circle has developed in which the more the outlet flow speed increases, the more the loss increases.
A desirable objective, therefore, has been of an overall three-dimensional design method that takes account of the effect by which the flow distribution in the circumferential direction is varied, and the effect of blade deformation due to centrifugal force. However, the prior art solutions to date have not eliminated all problems. One such solution now will be described with reference to FIGS. 14 and 15. A row of turbine moving blades is designed in a form in which the leading edge is twisted in the clockwise direction from the blade root to the blade tip. Therefore, when a tensile load due to centrifugal force acts on the blade effective portion 9, twist-return (untwisting ) occurs in the direction of arrow AR shown in FIG. 14. Accordingly, as shown in FIG. 15, the throatxc2x7pitch ratio (S/T) of the turbine moving blade 4, although set in the distribution shown by the solid line from blade root to blade tip when at rest, theoretically changes to the distribution shown by the broken line during operation. However, the measures taken to control vibration of the turbine moving blades (i.e., the intermediate connectors 11 in the intermediate part of the blade effective portion 9 and tip connectors 10 at the blade tips) restrict blade untwisting at these connecting points, and the throatxc2x7pitch ratio (S/T) distribution in the 70% to 95% height that is normalized between connectors 10 and 11, as shown in FIG. 15, swells outward and becomes a broad passage.
Further problems can result from this situation. In the case of the long blade turbine moving blades 4, where the diameter of the blade root is 1.4 m or more and the blade effective portion 9 exceeds 1 m, the equivalent speed of the motive steam leaving the turbine moving blade (the speed defined by coordinates set by the turbine moving blades) exceeds the speed of sound at least in the region from the mean diameter of the blade effective portion 9 (PCD: pitch circle diameter) to the blade tip, and becomes a supersonic speed flow. Given the range of turbine driving steam inlet flow angles, as shown in FIG. 13, along the blade effective portion 9, if the above-mentioned swollen portion occurs in the throatxc2x7pitch ratio (S/T) distribution in the 75% to 95% normalized blade height region, the supersonic flow of turbine driving steam will be expanded excessively, and a strong shock wave will be generated on the turbine moving blade.
Prior art steam turbines thus suffer from many drawbacks. They adopt throatxc2x7pitch ratio (S/T) distributions that yield almost uniform flow distributions in the radial direction, resulting in high frictional losses close to the wall surface at the blade roots of the turbine moving blades and close to the outer wall surface of the turbine nozzle blade tips. They also can suffer from shock waves caused by the interaction of supersonic steam flow with swollen blade portions between the restricted parts of the blade effective portion 9 due to blade untwisting. These drawbacks prevent the turbine from performing in accordance with design criteria.
It is an object of this invention to provide a steam turbine designed to improve turbine blade row performance.
It is a further object of this invention to provide a turbine moving blade which will make turbine driving steam flow in a stable state, thereby improving the performance of the turbine.
It is still a further object of this invention to provide a turbine nozzle blade which will make the turbine driving steam flow in a stable state, thereby improving performance of the turbine.
To achieve the objects, a three-dimensional blade design method devised and adopted for a turbine moving blade of the present invention is one that treats the turbine driving steam as a three-dimensional flow, and can control that three-dimensional flow. Therefore, accuracy is greater than with the prior art simplified three-dimensional blade design method.
Stated otherwise, in the turbine blade row, the throatxc2x7pitch ratio (S/T) of the turbine moving blades is off-set prior to operation. When blade untwist occurs during operation, excessive expanded flow in the supersonic speed region is prevented by producing an appropriate throatxc2x7pitch ratio (S/T) distribution corresponding to the turbine driving steam entry angle by maintaining proper values.
At the same time, a flow distribution is given in the radial direction so that, with both turbine moving blades and turbine nozzle blades, the turbine driving steam flow is reduced in regions close to the wall surface where losses otherwise would be large while, on the other hand, the turbine driving steam flow is increased in regions distant from the wall surface where losses are small.